View metadata, citation and similar papers at core.ac.uk brought to you by CORE provided by AUT Scholarly Commons Mathematical modeling of a solar powered humidification dehumidification desalination prototype Reza Enayatollahi1, Timothy Anderson1, Roy Nates1 1 Department of Mechanical Engineering, Auckland University of Technology, Auckland, New Zealand [email protected] Abstract In recent times the issue of fresh water shortages and salinity contamination of existing water sources has become a serious problem in a number of locations around the world. Hence, developing an environmentally friendly desalination technique is essential. In this work a theoretical model is developed in order to optimize a novel humidification-dehumidification desalination system. A sensitivity analysis was carried out, in order to find the optimum values for air and water flow rates. From this analysis it was found that a maximum of production rate of 1.5 kg/hr.m2 was achievable, however it was also found that this rate was particularly influenced by the incident radiation, the inlet water temperature and the water flow rate. Keywords: Humidification, Dehumidification, HDH, Desalination 1. Introduction Fresh water demand is increasing considerably with population growth and improvement of living standards. Currently more than half of world’s population is suffering from fresh water shortages [1]. Although nearly three-quarters of earth’s surface is covered with water, only 3% of this water is potable [2] that is to say, it contains less than 500 ppm of salt [3]. Many current desalination plants use fossil fuels as the source of energy, but these are costly and not environmentally friendly [4]. Therefore, a large number of investigations have been aimed at finding efficient and clean techniques for water refinement [5,6,7,8,9]. Solar energy has the potential to solve these problems. One of the cheapest and simplest ways to produce drinkable water is through the use of the Humidification Dehumidification (HDH) technique [10,11]. The HDH Desalination is a low temperature process often used to produce small amounts of desalted water [2]. Therefore it has the capability of using solar energy to drive it. The challenges for these systems are to minimize their size and complexity. Al-Hallaj et al, [10] designed and assessed a HDH Desalination unit with an open-air cycle and a capacity of 12 liter/day.m2 of solar absorber area. It was reported that for this system that there was an optimum airflow rate, above and below which the fresh water production rate was reduced. Zamen et al, [12] carried out an experimental investigation on a two-stage desalination unit with 80m2 of solar panels. According to their report, productivity of the entire unit could be increased by 20% when compared with a single stage unit. It was also observed that using a two-stage method could increase heat Solar2014: The 52nd Annual Conference of the Australian Solar Council recovery in the condenser and consequently decrease the size of the thermal energy system per unit of fresh water produced. Bacha et al, [13] conducted a mathematical and experimental study on a HDH Desalination unit in which both water and air heating techniques were applied. According to their work the two-temperature mathematical model better describes the real behavior of the water collector compared to the one- temperature model. Given the potential for HDH desalination, the aim of this work is to develop a generalized design model to aid the development of such systems and predict their performance and optimal operating conditions. 2. System Description The present study is based on the open air-closed water HDH cycle, for a small household, as shown in Figure 1. For the proposed system a double-glazed flat plate solar collector with an absorber area of 2 m2 is used for heating the saline water. The water from a very large reservoir, which is fed by the preheated water from the condenser together with make-up water from the saline source, enters the collector where it is heated, providing the required heat for evaporation, before exiting into the evaporator. The evaporator acts as the humidifier and is of a cascading design. In doing this, it utilizes a counter-current flow arrangement with ambient air introduced into the system. Figure 1. HDH Desalination Unit As a result of the contact between the heated water and the air, some water will be evaporated, ideally saturating the air. This evaporation will cause a temperature drop in the remaining water, and if the evaporation area is large enough relative to the water flow rate, the outgoing water from the evaporator can be used as the cooling fluid in the condenser. This cooled water exiting the evaporator, enters the condenser, which consists of a shell and tube heat exchanger. Cold water from the evaporator enters on the shell side and hot moist air enters the tubes. As a result of heat transfer between these two streams Solar2014: The 52nd Annual Conference of the Australian Solar Council water vapor in the air will be condensed, thus delivering potable water. The design parameters for the proposed system are shown in Table 1. Table 1. Design Parameters Water collector Evaporator Condenser Ambient Condition 293.2 294.2 20 60 T T N Rh , , [K] , , [K] , , [%] -5 w in col 1.17*10 a in sys 0.017 tube cond 2 a amb 38 V 3 V , , 3 L s , , [m /s] [m /s] [m] [°] ̇w in col 20 ̇a in sys 29 cond 19 1000 N , N , N , G 2 [W/m ] tube col 0.0127 Tray evap 2 baffle cond 0.0127 t 101.325 D L D P , , [mm] [m] , , [mm] [kpa] o tube col 0.01181 evap 1 o tube cond 0.01181 sys 5 D , , W D , , V [mm] [m] [mm] [m/s] i tube col 1 evap 0.4 i tube cond 0.2129 wind 0.4615 N , D D , R [m] [m] [kJ/kg.K] glass col evap shell cond v The water reservoir is considered sufficiently large so that the inlet water temperature can be assumed constant. It is decided that the power required for the fan and pump would not be a concern of this investigation, but would be provided by an auxiliary PV system. 3. Method As mentioned earlier a standard double-glazed, flat plate solar collector is designed in order to heat up the water before entering the evaporator. The outlet water temperature from the collector can be calculated from Equation 1. , , , = , , + (1) , 푄, 푢 푐표푙 , 푤 표푢푡 푐표푙 푤 푖푛 푐표푙 푤 푖푛 푐표푙 푤 푐표푙 푇 In which,푇 , 푚, ̇ , ,푐푝 and , , are the specific heat, mass flow rate and temperature of inlet water to the collector, respectively. represents the 푤 푐표푙 푤 푖푛 푐표푙 푤 푖푛 푐표푙 , useful energy 푐푝gained 푚bẏ water and푇 it is determined by solving the following 푢 푐표푙 Equations 2 and 3, simultaneously. 푄 , = [ , ] (2) 푢 푐표푙 푐표푙 푎 ,푐표푙 푝 푐표푙 푎푚푏 푄 , = 퐴, , 퐺 +− 푈 �푇 (1− 푇 , �) (3) 푄푢 푐표푙⁄퐴. 푐표푙 푝 푐표푙 푤 푖푛 푐표푙 푅 푐표푙 푇 where,푇 is푈 the푐표푙퐹푅 collector푐표푙 − 퐹 overall heat loss coefficient , is the absorber plate temperature푐표푙 , is the collector area, is the absorbed푝 푐표푙 radiation, is the ambient 푈temperature and , is the heat removal factor푇 calculated by 푐표푙 푎 푎푚푏 Equation 4. 퐴 퐺 푇 퐹푅 푐표푙 , , , , , ,́ , = 1 −퐴푐표푙푈푐표푙퐹푐표푙 (4) 푚̇ 푤 푖푛 푐표푙푐푝푤 푐표푙 푚̇ 푤 푖푛 푐표푙푐푝푤 푐표푙 푅 푐표푙 퐹 In Equation퐴푐표푙푈푐표푙 (4), � − 푒 is the collector� efficiency factor defined by Equation 5. 퐹′푐표푙 Solar2014: The 52nd Annual Conference of the Australian Solar Council = (5) 1⁄푈푐표푙 , , , , 푐표푙 1 , , 1 1 퐹′ 푤� +퐶 +ℎ 휋퐷 � Where푈푐표푙, �2퐿 푠is퐹+ tube퐷표 푡푢푏푒 spacing,푐표푙� 퐵 푖 푡푢푏푒 푐표푙is the푖 푡푢푏푒 bond푐표푙 conductance, is the tube pitch, is the fin efficiency given by Equation 6 and , , is the heat transfer 퐵 푠 coefficient of푤 water inside the collector's퐶 tubes. Finally, 퐿 and 푖 푡푢푏푒 푐표푙 , , , , are퐹 the inner and outer diameters of the collectors tubes,ℎ respectively. 푖 푡푢푏푒 푐표푙 표 푡푢푏푒 푐표푙 ( ) ( ) 퐷 퐷 = ( ) ( ) (6) 1 exp 푚퐿푠 −exp −푚퐿푠 퐹 푚Where,퐿푠 �exp 푚 퐿푠 +exp −푚퐿푠 � = 푚 � where,푈푐표푙⁄푘 푝푙푝 and are thermal conductivity and thickness of absorber plate. The collector푘푝 overall푙푝 heat loss coefficient is the summation of bottom, edge and top heat loss coefficients. Bottom and edge heat loss coefficient can be 푐표푙 calculated from equations 7 and 8, respectively. 푈 = (7) 푘푖푛푠 푏표푡푡표푚 푙푖푛푠 푈 = (8) 퐴푝푒푟푖푚푒푡푒푟 푈푒푑푔푒In which,푈′푒푑푔푒 � 퐴푐표푙 and� are the thermal conductivity and thickness of insulation. is the edge overall loss coefficient and was assumed to be 0.5 푘푖푛푠 푙푖푛푠 for design purposes푒푑푔푒 , and is the edge area. In order to calculate the top loss coefficient푈′ , Equation 9 can be used [14]. 퐴푝푒푟푖푚푒푡푒푟 , −1 ( , )( , ) = . + + (9) , 2 , 2 푁푔 푐표푙 1 [ , .휎 푇푝, 푐표푙+푇푎푚푏, 푇]푝 푐표푙 +푇푎푚푏 , 0 33 2푁 , +퐵−1 푡표푝 , 푇푝 푐표푙, −푇 푤푖푛푑 −1 푔 푐표푙 푈 � 퐺 푎푚푏 ℎ � 푝 푐표푙 푔 푐표푙 푝 푐표푙 푔 푐표푙 휀 +0 5푁 �1−휀 � + 휀푔 푐표푙 −푁 푇푝 푐표푙� 푁푔 푐표푙+퐵 � where , is the number of glass covers, is the total radiation, and are empirical variables, is the Stephan-Boltzman constant and , and , are 푁푔 푐표푙 emissivity of absorber plate and glass cover. 퐺 퐵 푍 휎 휀푝 푐표푙 휀푔 푐표푙 = 250[1 0.0044( 90)] 푍 = 1 0.−04 +푠0−.0005 [1 + 0.091 , ] 2 퐵 �Where− ℎ푤푖푛푑 is wind heatℎ푤푖푛푑 transfer� coefficient푁푔 푐표푙 and is the slope of water collector. ℎ푤푖푛푑 푠 The evaporator is modeled considering the counter-current flow pattern for water and air streams.